U.S. patent application number 12/235084 was filed with the patent office on 2009-05-14 for method and apparatus of fuelling an internal combustion engine with hydrogen and methane.
Invention is credited to William Kendal Bushe, Gordon P. McTaggart-Cowen, Sandeep Munshi, Steven N. Rogak.
Application Number | 20090120385 12/235084 |
Document ID | / |
Family ID | 36585936 |
Filed Date | 2009-05-14 |
United States Patent
Application |
20090120385 |
Kind Code |
A1 |
Munshi; Sandeep ; et
al. |
May 14, 2009 |
Method And Apparatus Of Fuelling An Internal Combustion Engine With
Hydrogen And Methane
Abstract
A gaseous-fuelled internal combustion engine and a method of
engine operation improve combustion stability and reducing
emissions of NOx, PM, and unburned hydrocarbons. The method
comprises fuelling an internal combustion engine with hydrogen and
natural gas, which can be directly injected into the combustion
chamber together or introduced separately. Of the total gaseous
fuel delivered to the engine, at least 5% by volume at standard
temperature and pressure is hydrogen. For at least one engine
operating condition, the ratio of fuel rail pressure to peak
in-cylinder pressure is at least 1.5:1. A fuel injection valve
introduces the gaseous fuel mixture directly into the combustion
chamber. Two separate fuel injection valves could also introduce
the methane and hydrogen separately. An electronic controller
controls timing for operating the fuel injection valve(s). The
engine has a preferred compression ratio of at least 14:1.
Inventors: |
Munshi; Sandeep; (Vancouver,
CA) ; McTaggart-Cowen; Gordon P.; (Derby, GB)
; Rogak; Steven N.; (Vancouver, CA) ; Bushe;
William Kendal; (Vancouver, CA) |
Correspondence
Address: |
MCANDREWS HELD & MALLOY, LTD
500 WEST MADISON STREET, SUITE 3400
CHICAGO
IL
60661
US
|
Family ID: |
36585936 |
Appl. No.: |
12/235084 |
Filed: |
September 22, 2008 |
Related U.S. Patent Documents
|
|
|
|
|
|
Application
Number |
Filing Date |
Patent Number |
|
|
PCT/CA2007/000431 |
Mar 13, 2007 |
|
|
|
12235084 |
|
|
|
|
Current U.S.
Class: |
123/3 ;
123/575 |
Current CPC
Class: |
F02M 21/0215 20130101;
Y02T 10/36 20130101; F02B 69/04 20130101; Y02T 10/12 20130101; F02B
43/10 20130101; F02D 19/10 20130101; F02B 23/0657 20130101; F02B
2275/16 20130101; F02B 2275/14 20130101; F02M 21/0224 20130101;
Y02T 10/30 20130101; F02D 19/022 20130101; F02D 19/0605 20130101;
F02M 2200/46 20130101; F02B 3/06 20130101; Y02T 10/123 20130101;
F02D 19/0689 20130101; F02M 26/23 20160201; F02B 29/0468 20130101;
F02D 19/081 20130101; Y02T 10/32 20130101; F02B 23/0696 20130101;
F02M 43/04 20130101; Y02T 10/121 20130101; F02D 19/024 20130101;
F02M 25/10 20130101; C10L 3/06 20130101; F02D 19/0644 20130101;
F02B 37/00 20130101 |
Class at
Publication: |
123/3 ;
123/575 |
International
Class: |
F02B 43/10 20060101
F02B043/10 |
Foreign Application Data
Date |
Code |
Application Number |
Mar 31, 2006 |
CA |
2,539,711 |
Claims
1. A method of operating a direct injection internal combustion
engine, the method comprising introducing a gaseous fuel mixture
directly into a combustion chamber of said engine, wherein said
gaseous fuel mixture comprises methane and between 5% and 60%
hydrogen by volume at standard temperature and pressure, and for at
least one engine operating condition, maintaining a fuel rail to
peak in-cylinder pressure ratio of at least 1.5:1 when introducing
the gaseous fuel mixture into said combustion chamber.
2. The method of claim 1 wherein said gaseous fuel mixture
comprises between 10% and 50% hydrogen by volume at standard
temperature and pressure.
3. The method of claim 1 wherein said gaseous fuel mixture
comprises between 15% and 40% hydrogen by volume at standard
temperature and pressure.
4. The method of claim 1 wherein said gaseous fuel mixture
comprises between 20% and 35% hydrogen by volume at standard
temperature and pressure.
5. The method of claim 1 wherein said methane is a constituent part
of natural gas.
6. The method of claim 1 further comprising premixing said gaseous
fuel mixture and storing it as a blended fuel within a storage tank
from which it can be delivered to said engine.
7. The method of claim 1 further comprising controlling fuel
injection timing so that the mid-point of integrated combustion
heat release occurs between 2 and 30 crank angle degrees after top
dead center.
8. The method of claim 1 further comprising controlling fuel
injection timing so that in at least one engine operating condition
the mid-point of integrated combustion heat release occurs between
5 and 15 crank angle degrees after top dead center.
9. The method of claim 1 further comprising injecting a pilot fuel
directly into said combustion chamber about 1 millisecond before
start of injection of said gaseous fuel mixture.
10. The method of claim 9 wherein said pilot fuel is a liquid fuel
with a cetane number between 40 and 70.
11. The method of claim 10 wherein said liquid fuel is diesel
fuel.
12. The method of claim 9 wherein over an engine operating map said
pilot fuel is on average between 3% and 10% of the fuel that is
consumed by said engine on an energy basis.
13. The method of claim 9 wherein over an engine operating map said
pilot fuel is on average between 4% and 6% of the fuel that is
consumed by said engine on an energy basis.
14. The method of claim 1 further comprising heating a hot surface
inside said combustion chamber to assist with igniting said gaseous
fuel mixture.
15. The method of claim 14 wherein said hot surface is provided by
a glow plug and said method further comprises electrically heating
said glow plug.
16. The method of claim 1 further comprising spark igniting said
gaseous fuel mixture inside said combustion chamber.
17. The method of claim 1 further comprising storing said hydrogen
separately from said methane and mixing said hydrogen and methane
to form said gaseous fuel mixture.
18. The method of claim 1 further comprising controlling the
proportions of hydrogen and methane in said gaseous fuel mixture as
a function of engine operating conditions.
19. The method of claim 1 further comprising maintaining a fuel
rail to peak in-cylinder pressure ratio of at least 1.5:1 when
introducing said gaseous fuel mixture into said combustion chamber
for all engine operating conditions.
20. The method of claim 1 further comprising maintaining a fuel
rail to peak in-cylinder pressure ratio of at least 2:1 when
introducing said gaseous fuel mixture into said combustion chamber
for at least one engine operating condition.
21. The method of claim 1 further comprising maintaining a choked
flow condition at a nozzle orifice of a fuel injection valve when
introducing said gaseous fuel mixture into said combustion
chamber.
22. The method of claim 1 further comprising injecting said gaseous
fuel mixture into said combustion chamber with an injection
pressure that is at least 16 MPa (about 2350 psia).
23. The method of claim 1 further comprising injecting said gaseous
fuel mixture into said combustion chamber with an injection
pressure that is at least 20 MPa (about 2900 psia).
24. The method of claim 1 wherein in the course of a compression
stroke, an intake charge inside said combustion chamber is
compressed by a ratio of at least about 14:1.
25. The method of claim 1 wherein methane is the largest
constituent of said gaseous fuel mixture by volume at standard
temperature and pressure.
26. A method of fuelling an internal combustion engine, the method
comprising: introducing a gaseous fuel mixture directly into a
combustion chamber of said engine, wherein said gaseous fuel
mixture comprises methane; introducing hydrogen into said
combustion chamber, thereby adding hydrogen to said gaseous fuel
mixture, wherein said hydrogen represents between 5% and 60% by
volume of said gaseous fuel mixture at standard temperature and
pressure; and maintaining a gaseous fuel mixture rail to peak
in-cylinder pressure ratio of at least 1.5:1 when introducing the
gaseous fuel mixture into said combustion chamber for at least one
engine operating condition.
27. The method of claim 26 further comprising premixing said
hydrogen with said gaseous fuel mixture comprising methane, and
introducing said gaseous fuel mixture and said hydrogen directly
into said combustion chamber.
28. The method of claim 26 further comprising premixing said
hydrogen with intake air and introducing said hydrogen into said
combustion chamber during an intake stroke of said piston.
29. The method of claim 26 further comprising introducing said
hydrogen directly into said combustion chamber separately from said
gaseous fuel mixture.
30. An internal combustion engine capable of being fuelled with a
gaseous fuel mixture comprising methane and between 5% and 60%
hydrogen by volume at standard temperature and pressure, the engine
comprising: a combustion chamber defined by a cylinder, a cylinder
head, and a piston movable within said cylinder; a fuel injection
valve with a nozzle that is disposed within said combustion
chamber, said fuel injection valve being operable to introduce said
gaseous fuel mixture directly into said combustion chamber; a
pressurizing device and piping for delivering said gaseous fuel
mixture to said injection valve with a ratio of fuel rail to peak
in-cylinder pressure being at least 1.5:1 for at least one engine
operating condition; and an electronic controller in communication
with an actuator for said fuel injection valve for controlling
timing for operating said fuel injection valve.
31. The engine of claim 30 wherein said engine has a compression
ratio of at least 14:1.
32. The engine of claim 30 wherein said electronic controller is
programmable to time introduction of said gaseous fuel mixture into
said combustion chamber so that the mid-point of an integrated
combustion heat release occurs between 2 and 30 crank angle degrees
after top dead center.
33. The engine of claim 30 wherein said electronic controller is
programmable to time introduction of said gaseous fuel mixture into
said combustion chamber so that the mid-point of an integrated
combustion heat release occurs between 5 and 15 crank angle degrees
after top dead center.
34. The engine of claim 30 wherein said fuel injection valve is
mounted in said cylinder head with said fuel injection valve
comprising a nozzle disposed within said combustion chamber.
35. The engine of claim 30 further comprising a second fuel
injection valve that is operable to introduce a pilot fuel directly
into said combustion chamber.
36. The engine of claim 35 wherein said second fuel injection valve
is integrated into a valve assembly that also comprises said fuel
injection valve for introducing said gaseous fuel mixture.
37. The engine of claim 36 wherein said second fuel injection valve
and said fuel injection valve for introducing said gaseous fuel
mixture can be independently actuated and said gaseous fuel mixture
is injectable into said combustion chamber through a first set of
nozzle orifices, which are different from a second set of nozzle
orifices through which said pilot fuel is injectable into said
combustion chamber.
38. The engine of claim 30 further comprising an ignition plug
disposed within said combustion chamber that is operable to assist
with ignition of the gaseous fuel mixture.
39. The engine of claim 38 wherein said ignition plug is a glow
plug that is electrically heatable to provide a hot surface for
assisting with ignition of said gaseous fuel mixture.
40. The engine of claim 38 wherein said ignition plug is a spark
plug.
41. The engine of claim 30 further comprising a storage vessel for
storing said gaseous fuel mixture in a substantially homogeneous
mixture with predetermined proportions of hydrogen and methane.
42. The engine of claim 30 further comprising a first storage
vessel within which said hydrogen can be stored, a second storage
vessel within which a gaseous fuel comprising methane can be
stored, and valves associated with each one of said first and
second storage vessel that are operable to control respective
proportions of hydrogen and methane in said gaseous fuel mixture
that is introducible into said combustion chamber.
43. The engine of claim 42 wherein said electronic controller is
programmable to change respective proportions of hydrogen and
methane in said gaseous fuel mixture to predetermined amounts
responsive to detected engine operating conditions.
44. An internal combustion engine capable of being fuelled with a
gaseous fuel mixture comprising methane and hydrogen, the engine
comprising: a combustion chamber defined by a cylinder, a cylinder
head, and a piston movable within said cylinder; a first fuel
injection valve with a nozzle disposed within said combustion
chamber, wherein said fuel injection valve is operable to introduce
methane directly into said combustion chamber; a second fuel
injection valve with a nozzle disposed within an intake air
manifold, wherein said second fuel injection valve is operable to
introduce hydrogen into said intake air manifold from which said
hydrogen can flow into said combustion chamber; and an electronic
controller electrically connected to an actuator for each one of
said first and second fuel injection valves for controlling
respective timing for operating said first and second fuel
injection valves.
45. The engine of claim 44 further comprising a pressurizing device
and piping for delivering said methane to said first injection
valve with a ratio of fuel rail pressure to peak in-cylinder
pressure being at least 1.5:1 for at least one engine operating
condition.
46. The engine of claim 44 wherein said engine has a compression
ratio of at least 14:1.
Description
CROSS-REFERENCE TO RELATED APPLICATION(S)
[0001] This application is a continuation of International
Application No. PCT/CA2007/000431, having an international filing
date of Mar. 13, 2007, entitled "Method And Apparatus Of Fuelling
An Internal Combustion Engine With Hydrogen And Methane". The '431
international application claimed priority benefits, in turn, from
Canadian Patent Application No. 2,539,711 filed Mar. 31, 2006. The
'431 international application is hereby incorporated by reference
herein in its entirety.
FIELD OF THE INVENTION
[0002] The present invention relates to a method and apparatus of
fuelling a diesel-cycle internal combustion engine with hydrogen
and methane to improve combustion stability and reduce emissions of
nitrogen oxides (NOx), unburned hydrocarbons and particulate matter
(PM).
BACKGROUND OF THE INVENTION
[0003] Because gaseous fuels such as natural gas, propane,
hydrogen, and blends thereof are cleaner burning fuels compared to
liquid fuels such as diesel, recent attention has been directed to
developing engines that can burn such fuels while matching the
power and performance that engine operators are accustomed to
expecting from diesel engines.
[0004] Natural gas fuelled engines that use lean-burn
spark-ignition ("LBSI") introduce the fuel into the intake air
manifold or intake ports at relatively low pressures. To avoid
engine knock caused by the premature detonation of the fuel inside
the combustion chamber, such engines typically operate with a
compression ratio no greater than about 12:1, which is lower
compared to diesel-cycle engines which have compression ratios of
at least 14:1, and this affects engine performance and efficiency.
Consequently, while the exhaust gases from the combustion chambers
of LBSI engines can have lower emissions of NOx, and PM compared to
an equivalently sized diesel engine, such LBSI engines also have
lower performance and energy efficiency, which means that to do the
same amount of work, more fuel is consumed on an energy basis, and
to match the full range of power and performance of a diesel
engine, a larger LBSI engine is needed.
[0005] Recently, research has been directed towards blending
natural gas and hydrogen for use in homogeneous charge,
spark-ignition engines. Representative publications relating to
such research include, "The Effects of Hydrogen Addition On Natural
Gas Engine Operation", SAE Technical Paper 932775, by M. R. Swain,
M. J. Yusuf, Z. Dulger and M. N. Swain, which was published by the
Society of Automotive Engineers ("SAE") in 1993; "Variable
Composition Hydrogen/Natural Gas Mixtures for Increased Engine
Efficiency and Decreased Emissions", ASME Journal of Engineering
for Gas Turbines and Power, Vol. 122, pp. 135-140, by R. Sierens
and E. Rousseel, published in 2000; "Hydrogen Blended Natural Gas
Operation of a Heavy Duty Turbocharged Lean Burn Spark Ignition
Engine", SAE Technical Paper 2004-01-2956, by S. R. Munshi, C.
Nedelcu, J. Harris, et al., published in 2004; "Hydrogen
Enrichment: A Way to Maintain Combustion Stability in a Natural Gas
Fuelled Engine with Exhaust Gas Recirculation, the Potential of
Fuel Reforming", Proceedings of the Institution of Mechanical
Engineers, Part D. Vol. 215 2001, pp. 405-418, by S. Allenby, W-C.
Chang, A. Megaritis and M. L. Wyszynski; "Emission Results from the
New Development of a Dedicated Hydrogen-Enriched Natural Gas
Heavy-Duty Engine", SAE Technical Paper 2005-010235, by K. Collier,
N. Mulligan, D. Shin, and S. Brandon which was published in 2005;
"Comparisons of Emissions and Efficiency of a Turbocharged
Lean-Burn Natural Gas and Hythane-Fuelled Engine", ASME Journal of
Engineering for Gas Turbines and Power, Vol. 119, 1997, pp.
218-226, by J. F. Larsen and J. S. Wallace; "Effect of hydrogen
addition on the performance of methane-fuelled vehicles. Part I:
effect on S.I. engine performance", International Journal of
Hydrogen Energy, Vol. 26. 2001, pp. 55-70, by CG. Bauer and T. W.
Forest; "Methane-Hydrogen Mixtures as Fuels", International Journal
of Hydrogen Energy, Vol. 21 No. 7, 1996, pp. 625-631, by G. A.
Karim, I. Wierzba and Y. Al-Alousi; and "Internal Combustion
Engines Fuelled by Natural Gas-Hydrogen Mixtures", International
Journal of Hydrogen Energy, Vol. 29, 2004, pp. 1527-1539, by S. O.
Akansu, Z. Dulger, N. Kahraman and T. Veziroglu. The results
reported in these papers have shown that at stoichiometric
operation, the addition of hydrogen tends to reduce power density
and increase NOx, while slightly reducing hydrocarbon and carbon
monoxide emissions. A more significant effect is reported under
lean premixed conditions, where a substantial increase in the lean
limit is observed. This has been attributed to enhanced combustion
rate and shorter ignition delay. For a given air-fuel ratio, NOx
emissions are higher with hydrogen addition, due to the higher
flame temperature, while CO and unburned hydrocarbons are
substantially reduced. However, due to hydrogen's ability to extend
the lean limit, lower NOx emissions can be achieved by running at
leaner air-fuel ratios with hydrogen addition. Flame stability in
the presence of exhaust gas recirculation (EGR) is also improved.
Efficiency effects can depend upon the tested operating condition,
with some studies such as those reported in the Swain, Sierens, and
Akansu papers, showing improved efficiency with hydrogen addition
and other studies, such as those reported in the Larsen and Bauer
papers, showing reduced efficiency. Such contradictory results show
that while a considerable amount of research has been done to
investigate the effects of blending natural gas and hydrogen for
use in homogeneous charge spark-ignition engines, the combustion
process is complex, that the effect of combusting such fuel
mixtures in an engine can be very dependent upon the engine
operating conditions, and that the effect of adding hydrogen and
the magnitude or such effects, if any, are not obvious or easy to
predict. Furthermore, all of the published papers referenced herein
relate to homogeneous charge spark-ignition engines, and while some
laboratory experiments have been reported, such as shock-tube
studies and non-premixed counterflow methane/heated air jet
experiments, the inventors are not aware of any publications
relating to experiments involving fuelling a direct injection
internal combustion engine with a blended fuel mixture comprising
methane and hydrogen.
[0006] Engines that are capable of injecting a gaseous fuel
directly into the combustion chamber of a high compression internal
combustion engine are being developed, but are not yet commercially
available. Engines fuelled with natural gas that use this approach
can substantially match the power, performance and efficiency
characteristics of a diesel engine, but with lower emissions of
NOx, unburned hydrocarbons, and PM. NOx are key components in the
formation of photochemical smog, as well as being a contributor to
acid rain. PM emissions, among other detrimental health effects,
have been linked to increased cardiovascular mortality rates and
impaired lung development in children. However, with direct
injection engines that are fuelled with natural gas, it has been
found that there is a trade-off between NOx emissions and emissions
of unburned hydrocarbons and PM. That is, later timing for
injecting the natural gas is beneficial for reducing NOx but
results in higher emissions of unburned hydrocarbons and PM.
Environmental regulatory bodies in North America and around the
world have legislated substantial reductions in NOx and PM
emissions from internal combustion engines. As a result, because it
is necessary to reduce the emissions of each one of NOx, PM and
unburned hydrocarbons, for a direct injection engine fuelled with
natural gas, the higher PM emissions associated with later
combustion timing effectively limits how much the timing for fuel
injection can be retarded.
[0007] Since published technical papers have reported that under
specific operating conditions there can be benefits arising from
fuelling a homogeneous charge, spark-ignition engine with a gaseous
fuel mixture comprising methane and hydrogen, and since
environmental regulatory bodies have legislated substantial
reductions in NOx and PM emissions from internal combustion
engines, and since the combustion process is complex and the effect
of adding hydrogen to a fuel mixture delivered to a direct
injection internal combustion engine is unpredictable, there is a
need to determine whether it is possible to improve combustion
stability and reduce engine emissions by fuelling a direct
injection internal combustion engine with hydrogen and natural gas,
and if so, the method of operating a direct injection engine that
is fuelled with such fuels to achieve improvements in combustion
stability and reductions in engine emissions.
SUMMARY OF THE INVENTION
[0008] A method of operating a direct injection internal combustion
engine comprises introducing a gaseous fuel mixture directly into a
combustion chamber of the engine. The gaseous fuel mixture
comprises methane and between 5% and 60% hydrogen by volume at
standard temperature and pressure. For at least one engine
operating condition, the method comprises maintaining a fuel rail
to peak in-cylinder pressure ratio of at least 1.5:1 when
introducing the gaseous fuel mixture into the combustion chamber. A
preferred embodiment of the method comprises maintaining a fuel
rail to peak in-cylinder pressure ratio of at least 1.5:1 when
introducing the gaseous fuel mixture into the combustion chamber
for all engine operating conditions. When the constituent parts of
the gaseous fuel mixture are described herein as percentages by
volume, unless noted otherwise this is defined to be the percentage
by volume at standard temperature and pressure (STP).
[0009] In preferred methods, the gaseous fuel mixture can comprise
between 10% and 50%, between 15% and 40% hydrogen by volume, or
between 20% and 35% hydrogen by volume at standard temperature and
pressure. The methane can be a constituent part of natural gas. The
method can further comprise premixing the gaseous fuel mixture and
storing it as a blended fuel within a storage tank from which it
can be delivered to the engine. In a preferred method, methane is
the largest constituent of the gaseous fuel mixture by volume at
standard temperature and pressure.
[0010] The method can further comprise controlling fuel injection
timing so that the mid-point of integrated combustion heat release
occurs between 2 and 30 crank angle degrees after top dead center.
An advantage of adding hydrogen to natural gas is that the
combustion timing can be delayed to a later time in the combustion
cycle compared to an engine that is fuelled with natural gas alone.
A preferred method comprises controlling fuel injection timing so
that in at least one engine operating condition the mid-point of
integrated combustion heat release occurs between 5 and 15 crank
angle degrees after top dead center.
[0011] The method can comprise introducing a pilot fuel to assist
with ignition of the gaseous fuel mixture. A preferred method
comprises injecting a pilot fuel directly into the combustion
chamber about 1 millisecond before start of injection of the
gaseous fuel mixture. The pilot fuel can be a liquid fuel with a
cetane number between 40 and 70. A pilot fuel with a cetane number
between 40 and 50 is preferred in most cases, with conventional
road grade diesel being a suitable fuel with a cetane number in
this range. Over an engine operating map the pilot fuel is on
average between 3% and 10% of the fuel that is consumed by the
engine on an energy basis, and more between 4% and 6%. The pilot
fuel is more easily ignited compared to the gaseous fuel mixture,
and the pilot fuel ignites first to trigger the ignition of the
gaseous fuel mixture. Because the gaseous fuel mixture is
preferably cleaner burner than the pilot fuel, the pilot fuel
preferably represents only a small portion of the fuel that is
consumed by the engine on an energy basis.
[0012] Instead of employing a pilot fuel, the method can comprise
heating a hot surface inside the combustion chamber to assist with
igniting the gaseous fuel mixture. In a preferred method the hot
surface is provided by a glow plug and the method further comprises
electrically heating the glow plug. In yet another embodiment, the
method can comprise spark igniting the gaseous fuel mixture inside
the combustion chamber.
[0013] The method can further comprise storing the hydrogen
separately from the methane and mixing the hydrogen and methane to
form the gaseous fuel mixture. The method can further comprise
controlling the proportions of hydrogen and methane in the gaseous
fuel mixture as a function of engine operating conditions.
[0014] The method can further comprise maintaining a fuel rail to
peak in-cylinder pressure ratio of at least 2:1 when introducing
the gaseous fuel mixture into the combustion chamber for at least
one engine operating condition. Preferred methods comprise
maintaining a choked flow condition at a nozzle orifice of a fuel
injection valve when introducing the gaseous fuel mixture into the
combustion chamber. While experiments have proven that satisfactory
engine operation can be achieved by injecting the gaseous fuel
mixture into the combustion chamber with an injection pressure that
is at least 16 MPa (about 2350 psia), higher fuel injection
pressures of at least 20 MPa (about 2900 psia) are more
preferred.
[0015] According to the method, in the course of a compression
stroke, an intake charge inside the combustion chamber is
compressed by a ratio of at least about 14:1. Compression ratios
higher than 14:1 are associated with diesel-cycle engines, which
can deliver higher performance and efficiency than conventional
Otto-cycle engines, otherwise known as spark-ignition engines,
which use a pre-mixed homogeneous charge which limits them to lower
compression ratios to avoid engine knock.
[0016] In another preferred method of fuelling an internal
combustion engine, the method comprises introducing a gaseous fuel
mixture directly into a combustion chamber of the engine, wherein
the gaseous fuel mixture comprises methane, introducing hydrogen
into the combustion chamber, thereby adding hydrogen to the gaseous
fuel mixture, wherein the hydrogen represents at least 5% by volume
of the gaseous fuel mixture at standard temperature and pressure;
and maintaining a gaseous fuel mixture rail to peak in-cylinder
pressure ratio of at least 1.5:1 when introducing the gaseous fuel
mixture into the combustion chamber for at least one engine
operating condition. That is, the hydrogen can be introduced into
the combustion chamber separately from the gaseous fuel mixture and
becoming part of the gaseous fuel mixture inside the combustion
chamber or the method can comprise premixing the hydrogen with the
gaseous fuel mixture comprising methane, and introducing the
gaseous fuel mixture and the hydrogen directly into the combustion
chamber. In further embodiments, the method can comprise premixing
the hydrogen with intake air and introducing the hydrogen into the
combustion chamber during an intake stroke of the piston or
introducing the hydrogen directly into the combustion chamber
separately from the gaseous fuel mixture.
[0017] An internal combustion engine is provided that can be
fuelled with a gaseous fuel mixture comprising methane and between
5% and 60% hydrogen by volume at standard temperature and pressure.
The disclosed engine comprises a combustion chamber defined by a
cylinder, a cylinder head, and a piston movable within the
cylinder; a fuel injection valve with a nozzle that is disposed
within the combustion chamber, the fuel injection valve being
operable to introduce the gaseous fuel mixture directly into the
combustion chamber; a pressurizing device and piping for delivering
the gaseous fuel mixture to the injection valve with a ratio of
fuel rail to peak in-cylinder pressure being at least 1.5:1 for at
least one engine operating condition; and, an electronic controller
in communication with an actuator for the fuel injection valve for
controlling timing for operating the fuel injection valve. The
engine preferably has a compression ratio of at least 14.
[0018] The electronic controller is preferably programmable to time
introduction of the gaseous fuel mixture into the combustion
chamber so that the mid-point of an integrated combustion heat
release occurs between 2 and 30 crank angle degrees after top dead
center, and in another embodiment, between 5 and 15 crank angle
degrees after top dead center.
[0019] The fuel injection valve can be mounted in the cylinder head
with the fuel injection valve comprising a nozzle disposed within
the combustion chamber. The engine can further comprise a second
fuel injection valve that is operable to introduce a pilot fuel
directly into the combustion chamber. The second fuel injection
valve can be integrated into a valve assembly that also comprises
the fuel injection valve for introducing the gaseous fuel mixture.
The second fuel injection valve and the fuel injection valve for
introducing the gaseous fuel mixture are preferably independently
actuated and the gaseous fuel mixture is injectable into the
combustion chamber through a first set of nozzle orifices, which
are different from a second set of nozzle orifices through which
the pilot fuel is injectable into the combustion chamber.
[0020] Instead of employing a second fuel injection valve to
introduce a pilot fuel to assist with ignition of the gaseous fuel
mixture, the engine can comprise an ignition plug disposed within
the combustion chamber that is operable to assist with ignition of
the gaseous fuel mixture. The ignition plug can be a glow plug that
is electrically heatable to provide a hot surface for assisting
with ignition of the gaseous fuel mixture or the ignition plug can
be a spark plug.
[0021] The engine can further comprise a storage vessel for storing
the gaseous fuel mixture in a substantially homogeneous mixture
with predetermined proportions of hydrogen and methane. In another
embodiment, the engine can comprise a first storage vessel within
which the hydrogen can be stored, a second storage vessel within
which a gaseous fuel comprising methane can be stored, and valves
associated with each one of the first and second storage vessel
that are operable to control respective proportions of hydrogen and
methane in the gaseous fuel mixture that is introducible into the
combustion chamber. If the hydrogen is stored separately from the
gaseous fuel mixture that comprises methane, then the electronic
controller can be programmable to change respective proportions of
hydrogen and methane in the gaseous fuel mixture to predetermined
amounts responsive to detected engine operating conditions.
[0022] Another embodiment of an internal combustion engine is
provided that can be fuelled with a gaseous fuel mixture comprising
methane and hydrogen. In this embodiment, the engine comprises a
combustion chamber defined by a cylinder, a cylinder head, and a
piston movable within the cylinder; a first fuel injection valve
with a nozzle disposed within the combustion chamber, wherein the
fuel injection valve is operable to introduce methane directly into
the combustion chamber; a second fuel injection valve with a nozzle
disposed within an intake air manifold, wherein the second fuel
injection valve is operable to introduce hydrogen into the intake
air manifold from which the hydrogen can flow into the combustion
chamber; and an electronic controller in communication with an
actuator for each one of the first and second fuel injection valves
for controlling respective timing for operating the first and
second fuel injection valves. In this embodiment, the engine can
further comprise a pressurizing device and piping for delivering
the methane to the first injection valve with a ratio of fuel rail
pressure to peak in-cylinder pressure being at least 1.5:1 for at
least one engine operating condition. Like the other embodiments,
the engine preferably has a compression ratio of at least 14:1, and
compression ratios as high as 25:1 are possible as well as ratios
therebetween, such as 18:1, 20:1 and 22:1.
BRIEF DESCRIPTION OF THE DRAWINGS
[0023] FIG. 1 is a schematic drawing illustrating an apparatus for
direct injection of a gaseous fuel mixture into the combustion
chamber of an internal combustion engine.
[0024] FIG. 2 is a schematic drawing illustrating a second
embodiment of an apparatus for direct injection of a gaseous fuel
mixture into the combustion chamber of an internal combustion
engine.
[0025] FIG. 3 shows four graphs that plot engine emissions against
timing for the mid-point of a combustion heat release for an engine
that is fuelled with 100% compressed natural gas, a gaseous fuel
mixture of 10% hydrogen and 90% compressed natural gas, and 23%
hydrogen and 77% compressed natural gas, with all percentages
measured by volume. The plotted data was collected from an engine
operating at 800 RPM, 6 bar GIMEP, 0.5.phi., 40% exhaust gas
recirculation (by mass), and with a fuel injection pressure of 16
MPa.
[0026] FIG. 4 shows four graphs that plot engine performance
characteristics against timing for the mid-point of a combustion
heat release for an engine that is fuelled with 100% compressed
natural gas, a gaseous fuel mixture of 10% hydrogen and 90%
compressed natural gas, and 23% hydrogen and 77% compressed natural
gas, with all percentages measured by volume. The engine operating
conditions were the same as for the data plotted in FIG. 3.
[0027] FIG. 5 shows two bar graphs that plot pilot and gaseous fuel
ignition delay for an engine that is fuelled with 100% compressed
natural gas, a gaseous fuel mixture of 10% hydrogen and 90%
compressed natural gas, and 23% hydrogen and 77% compressed natural
gas, with all percentages measured by volume. The engine operating
conditions were the same as for the data plotted in FIGS. 3 and
4.
[0028] FIG. 6 shows plots of in-cylinder pressure and heat release
rate for an engine that is fuelled with 100% compressed natural
gas, and a gaseous fuel mixture of 10% hydrogen and 90% compressed
natural gas with percentages for the gaseous fuel mixture measured
by volume and the timing for the mid-point of the integrated heat
release occurring at 10 crank angle degrees after top dead center.
The engine operating conditions were the same as for the data
plotted in FIGS. 3-5.
[0029] FIG. 7 is plots of in-cylinder pressure and heat release
rate for an engine that is fuelled with 100% compressed natural
gas, and a gaseous fuel mixture of 23% hydrogen and 77% compressed
natural gas, with percentages for the gaseous fuel mixture measured
by volume. In the upper two graphs, the timing for the mid-point of
the integrated heat release occurs at 5 crank angle degrees after
top dead center and in the lower two graphs the timing for the
mid-point of the integrated heat release occurs at 15 crank angle
degrees after top dead center. The engine operating conditions were
the same as for the data plotted in FIGS. 3-6.
[0030] FIG. 8 plots in-cylinder pressure and heat release rate for
constant and adjusted timing conditions for an engine fuelled with
100% compressed natural gas, and a gaseous fuel mixture of 23%
hydrogen and 77% compressed natural gas, with percentages for the
gaseous fuel mixture measured by volume. The engine operating
conditions were the same as for the data plotted in FIGS. 3-7.
[0031] FIG. 9 is a bar graph that plots gaseous fuel ignition delay
in crank angle degrees against timing for the mid-point of
integrated heat release for 100% natural gas, and a gaseous fuel
mixture of 23% hydrogen and 77% compressed natural gas, with
percentages for the gaseous fuel mixture measured by volume. The
plotted data shows the gaseous fuel ignition delay when the gaseous
fuel mixture is injected with the same timing that is employed when
the engine is fuelled with 100% natural gas, and the effect on
gaseous fuel ignition delay when timing is adjusted. The engine
operating conditions were the same as for the data plotted in FIGS.
3-8.
[0032] FIG. 10 plots engine emissions against the timing for the
mid-point of the integrated heat release, showing the effect of
increasing injection pressure from 16 MPa to 20 MPa for an engine
fuelled with 100% compressed natural gas, and a gaseous fuel
mixture of 23% hydrogen and 77% compressed natural gas, with
percentages for the gaseous fuel mixture measured by volume. Other
than the different fuel injection pressures, the engine operating
conditions were the same as for the data plotted in FIGS. 3-9.
[0033] FIG. 11 plots in-cylinder pressure and heat release rate
against crank angle degrees for fuel injection pressures of 16 MPa
and 20 MPa for an engine fuelled with 100% compressed natural gas,
and a gaseous fuel mixture of 23% hydrogen and 77% compressed
natural gas, with percentages for the gaseous fuel mixture measured
by volume. The engine operating conditions were the same as for the
data plotted in FIGS. 3-9.
DETAILED DESCRIPTION OF PREFERRED EMBODIMENT(S)
[0034] FIG. 1 is a schematic drawing of engine apparatus 100, which
is adapted to be fuelled with a gaseous fuel mixture comprising
methane and hydrogen. Fuel storage system 110 comprises storage
vessel 111, which is made to store the compressed gaseous fuel
mixture. In the illustrated embodiment, the pre-mixed gaseous fuel
mixture can be stored in storage vessel 111, which is rated to
store the compressed gaseous fuel mixture at a predetermined
storage pressure. Storage vessel 111 is designed to comply with
local regulations which can specify safety factors for ensuring
containment of the gaseous fuel mixture even in the event of
impact, for example if storage vessel 111 is a vehicular fuel tank
that could be involved in a vehicle collision. In addition to
safety factors and design strength requirements, local regulations
typically impose a maximum storage pressure. Compressor 112 is
operable to deliver the gaseous fuel mixture from storage vessel
111 to fuel injection valve 120, via aftercooler 113 and
gaseous-fuel supply rail 116. Aftercooler 113 cools the gaseous
fuel mixture after it has been compressed by compressor 112.
Aftercooler 113 can be a heat exchanger with the cooling fluid
being engine coolant diverted from the engine cooling system. In a
preferred embodiment, the fuel supply system is a common rail
system, meaning that the gaseous fuel is delivered to fuel
injection valve assembly 120 at injection pressure. In such a
common rail system, pressure sensor 115 can be employed to measure
the fuel pressure in gaseous-fuel supply rail 116 so that
compressor 112 can be operated to maintain gaseous fuel injection
pressure between a predetermined low and high set point.
[0035] In a preferred embodiment, a liquid pilot fuel is also
directly injected into combustion chamber 122 to assist with
igniting the gaseous fuel mixture. In such an embodiment, injection
valve assembly 120 can comprise two separate valve needles that are
independently operable, with one valve needle controlling the
injection of a gaseous fuel mixture and the second valve needle
controlling the injection of the liquid pilot fuel. Pilot fuel is
deliverable to fuel injection valve assembly 120 from pilot fuel
rail 118. Pilot fuel can be delivered to pilot fuel rail 118 at
injection pressure by a conventional diesel common rail fuel supply
system (not shown).
[0036] Fuel injection valve 120 introduces the gaseous fuel mixture
directly into combustion chamber 122, which is generally defined by
a bore provided in cylinder block 124, the cylinder head, and
piston 126, which is movable up and down within the bore. The flow
of air into combustion chamber 122 from intake air manifold 130 is
controlled by intake valve 132, which can be opened during intake
strokes of piston 126. Like conventional diesel engines, the
disclosed engine can employ a turbocharger (not shown) to
pressurize the intake air or the engine can be naturally aspirated.
Combustion products can be expelled from combustion chamber 122
into exhaust manifold 140 through exhaust valve 142, which can be
opened during exhaust strokes of piston 126.
[0037] Electronic controller 150 is programmable to control the
operation of compressor 112 and control valve 114 to control the
pressure of the gaseous fuel mixture in gaseous fuel supply rail
116. Controller 150 is also programmable to command the timing for
opening and closing of the fuel injection valve needles that
respectively control the injection of the gaseous fuel mixture and
the pilot fuel. For example, electronic controller 150 can be
programmed to control the pilot fuel injection valve so that the
pilot fuel is introduced about 1 millisecond before the gaseous
fuel injection valve is commanded to open. Furthermore, electronic
controller 150 can be programmed to time the opening and closing of
the gaseous fuel injection valve. The fuel injection timing can be
predetermined responsive to the engine operating conditions
determined from measured parameters that are inputted into
electronic controller 150, and the input of such parameters is
represented by arrow 152.
[0038] FIG. 2 is a schematic drawing of another preferred
embodiment for an engine apparatus that is adapted to be fuelled
with a gaseous fuel mixture comprising methane and hydrogen. In
this embodiment, fuel injection valve 220 injects only the gaseous
fuel mixture into combustion chamber 222. A pilot fuel is not
required by this engine because ignition assistance is provided by
ignition plug 228. Ignition plug 228 can be an electrically heated
glow plug that is adapted for sustained operation during engine
operation. This is unlike a conventional glow plug, which is
normally activated only under certain engine conditions such as
start-up when the engine block is below a predetermined
temperature. Compared to an engine that is fuelled with natural gas
without added hydrogen, an advantage of using a gaseous fuel
mixture comprising hydrogen is that because hydrogen is easier to
ignite compared to natural gas, the glow plug temperature can be
kept at a lower temperature compared to the temperature that is
needed to assist with ignition of natural gas which is not mixed
with hydrogen. This is advantageous because lower glow plug
temperatures are generally associated with improved durability and
longer service life. In yet another embodiment (not shown), the
ignition plug can be a spark plug.
[0039] In the illustrated embodiment of FIG. 2, other than using
ignition plug 228 to assist initiating fuel combustion instead of a
pilot fuel, the shown engine apparatus is essentially the same.
That is, a bore in cylinder block 224, cylinder head 225, and
piston 226, which is movable up and down within the cylinder bore,
all cooperate to define combustion chamber 222. Air can flow into
combustion chamber 222 through intake air manifold 230 when intake
valve 232 is open and electronic controller 250 is programmable to
control the timing for opening and closing fuel injection valve
220, and to control the temperature of ignition plug 228.
[0040] FIG. 2 also shows an optional secondary fuel injection valve
240 which can be employed to inject some of the gaseous fuel into
the intake air manifold. A port fuel injection valve is shown, but
a single fuel injection valve can be disposed further upstream in
the intake air manifold for introducing gaseous fuel into all of
the combustion chambers. Secondary fuel injection valve 240 can be
employed to introduce hydrogen into the combustion chamber
separately from a gaseous fuel mixture comprising methane, such as
natural gas. With such an embodiment, the methane and hydrogen
mixes inside the combustion chamber, but with the hydrogen more
evenly dispersed within the combustion chamber. An advantage of
this arrangement is that the hydrogen need not be compressed to as
high a pressure as it would need to be pressurized for direct
injection. Another advantage is that separately injecting the
methane and hydrogen allows the proportions of each fuel to be
adjusted for different engine operating conditions. However, a
disadvantage of this approach is that a secondary fuel injection
valve adds complexity and capital and maintenance costs to the
engine.
[0041] To determine a desired method of operating an internal
combustion engine with direct injection of a gaseous fuel mixture
comprising methane and hydrogen, experiments were conducted using a
single cylinder engine. The single cylinder engine was a
Cummins.TM. ISX series heavy-duty six-cylinder, four stroke, direct
injection diesel engine, modified to operate on only one cylinder.
The engine was further adapted for gaseous fuel operation using
Westport.TM. fuel injection and fuel supply systems. The engine
cylinder bore diameter was 137 millimeters, the piston stroke was
169 millimeters, and the displacement of the single cylinder was
about 2.5 liters. The connecting rod length was 261 millimeters and
the compression ratio was 17:1.
[0042] Because the experimental engine was a single cylinder
engine, the energy in the exhaust stream was too small to drive a
turbocharger to compress the intake air. To simulate the conditions
for a turbocharged engine, in the experiments an air compressor was
provided for the combustion air supply. The air compressor was
equipped with a refrigerated air dryer to remove water vapor (dew
point -4O.degree. C.) and filters to remove contaminants. The EGR
loop comprised an EGR cooler and a variable flow-control valve.
Maintaining the exhaust stream pressure approximately 10 kPa above
the intake pressure drove the recirculation of the exhaust gas.
[0043] The fuelling system provided gaseous fuel and diesel to the
engine's internal fuelling rails. The fuel injection valve was a
dual fuel injection valve operable to separately and independently
inject the gaseous fuel mixture as the main fuel, and diesel fuel
as the pilot fuel, with flow of the main and pilot fuels being
controlled by two concentric valve needles. Separate solenoid
actuated control valves were operable to control the actuation of
each valve needle to control the timing and duration of the
respective pilot and main fuel injection events. The pilot fuel
injection valve comprised a nozzle with 7 orifices, and the gaseous
fuel injection valve comprised a nozzle with 9 orifices, and the
injection angle was 18 degrees below the firedeck. Two separate
gaseous fuel supplies were used in the experiments. Commercial
natural gas (-96 mol % CH.sub.4, 2% C.sub.2H.sub.6, traces N.sub.2,
CO.sub.2, C.sub.3H.sub.8 all <0.5%) was compressed externally
and used as the reference gas. The hydrogen-methane blended gaseous
fuel mixtures were purchased separately (certified standard,
analytical accuracy .+-.2%), and were fed via a separate
compression system to the engine. Low-sulfur (<500 ppm) road
grade diesel that met Canadian General Standards Board
specification CAN/CGSB-3.520 was used as the pilot fuel.
[0044] The gas flow rate was measured by a coriolis-type mass
flowmeter (uncertainty of approximately 1% of full scale), which
was insensitive to the composition of the gas passing through it.
Diesel pilot flow was measured by a gravimetric system with an
uncertainty of about 10% of full scale. Air supply was measured by
a subsonic venturi with an estimated uncertainty of 3% of full
scale. Cylinder pressure was measured with a flush-mounted
water-cooled piezo-electric pressure transducer with an uncertainty
of 1%, and correlated with crank angle (CA) by a shaft encoder with
a Vi crank angle degree (.degree. CA) resolution. Gaseous emissions
were measured using a raw emissions bench equipped with infrared
analyzers (CO.sub.2-- Beckmann, CO and CH.sub.4--Siemens), a flame
ionization detector (total unburned hydrocarbons (tHC)--Ratfisch),
and a chemi-luminescent analyzer (NOx--Advanced Pollution
Instruments). A second infrared analyzer (California Analytical)
was used to measure the CO.sub.2 concentration in the intake
stream, from which the EGR fraction was determined. A chilled water
separator removed water vapor (dew point -5.degree. C.) upstream of
the non-dispersive infrared instruments. Repeatability studies on
the gaseous emissions sampling have shown uncertainties of 5% in
NOx and 10% in tHC and CO, including both instrumentation
uncertainty and variations in engine operating condition.
Particulate matter was measured using a micro-dilution system,
where a fraction of the exhaust stream was separated and diluted at
a factor of 15:1. The particulate loading in this diluted sample
was then measured either using a tapered element oscillating
microbalance ("TEOM"), Rupprecht & Pataschnick Model 1105, or
with gravimetric filters. Pallflex Emfab.TM. filters were used to
collect the samples, and were then weighed (accuracy .+-.5 .mu.g)
to calculate the mass concentration in the exhaust stream. TEOM
results were found to be, on average, 8% below the gravimetric
filter readings (correlation coefficient 0.96). Where TEOM results
are used in this work, they are identified by the caption "TEOM
PM".
[0045] Due to the single cylinder engine's high internal friction,
brake-performance parameters are not representative of the
in-cylinder conditions. As a result, the engine operation was
measured on the basis of the gross-indicated power--the integral of
the in-cylinder pressure versus volume curve, over the compression
and power strokes only, as defined in J. B. Heywood in "Internal
Combustion Engine Fundamentals, published in 1988 by McGraw-Hill,
New York. The gross-indicated power, normalized by engine speed and
displaced volume, provided the gross-indicated mean effective
pressure (GIMEP). The indicated power was used to normalize both
fuel consumption and emissions measurements. The gross-indicated
specific fuel consumption (GISFC) reported the total fuel mass
flow, with the gaseous component represented as an equivalent mass
of diesel on an energy basis (lower heating values: diesel, 42.8
MJ/kg; NG, 48.8 MJ/kg; 10% H.sub.2, 50.6 MJ/kg; 23% H.sub.2, 52.5
MJ/kg).
[0046] The in-cylinder pressure trace can also be used to estimate
the net heat-release rate, as given by:
Q net .theta. = .gamma. .gamma. - 1 p V .theta. + 1 .gamma. - 1 V p
.theta. ##EQU00001##
where .theta. is the crank angle, p is the in-cylinder pressure at
a given crank angle, V is the cylinder volume at that point, and
.gamma. is the specific heat ratio (c.sub.p/c.sub.v--assumed
constant). The net heat release rate represents the rate of energy
release from the combustion processes less wall heat transfer and
crevice flow losses. By integrating the heat-release rate up to a
given crank-angle and normalizing by the total energy released over
the full cycle, the fraction of the energy released up to that
point in the cycle can be determined. The midpoint of this curve is
50% of the integrated heat release (50% IHR), and can be used to
define the combustion timing.
[0047] The engine operation was also defined on the basis of the
equivalence ratio (.phi.: ratio of actual to stoichiometric
fuel/oxidizer ratio). The amount of dilution of the intake air is
defined by the intake oxygen mass fraction (Y.sub.intO2), which is
0.23 for undiluted air and decreases with increasing dilution (that
is, increasing EGR). By specifying .phi., 50% IHR, GIMEP, engine
speed and the intake oxygen mass fraction, the engine's operating
condition is fully defined.
[0048] The experimental test conditions selected for testing the
gaseous fuel mixture comprising methane and hydrogen were based on
a desire to reduce fuel consumption while increasing operating
condition realism. Specifically, an operating condition with high
emissions associated with natural gas operation was of interest, to
determine how effectively hydrogen could enhance poor natural gas
combustion. The selected operating condition had the following
characteristics: a high EGR fraction, namely 40% by mass; an intake
oxygen mass fraction (Y.sub.intO2) of 0.175; an engine speed of 800
RPM; a low load, namely 6 bars gross indicated mean effective
pressure ("GIMEP"); and, a moderate .phi. of 0.5 (oxygen-based).
Experiments were conducted with a fuel injection pressure of 16 MPa
and 20 MPa. Natural gas with a 94% methane concentration by volume
was the source of methane for the gaseous fuel mixture, and
mixtures with 10% hydrogen and 23% hydrogen by volume were tested.
To establish influences over a range of conditions while minimizing
the required changes to the operating condition, a range of
combustion timings were used. By varying combustion timing, highly
stable conditions (early timings) and very unstable conditions
(late timings) could be tested at the same baseline (EGR, load,
speed) condition. To improve experimental precision, it was decided
to use a paired-testing approach, where a single point was tested
using first natural gas and then the gaseous fuel mixture (or in
the opposite order). By fixing the operating condition, then
varying the timing, it was possible to minimize variations due to
non-repeatability of the operating condition setpoint. Replication
of timing sets was used to establish repeatability. Most of the
testing was carried out with a fuel injection pressure of 16 MPa,
to ensure that the commanded injection opening durations were
repeatable (in excess of 0.9 ms). As this pressure is below the
pressures typically used in other gaseous fuelled direct injection
internal combustion engines, such as engines that are fuelled with
100% natural gas, a set of tests, with both natural gas and the
gaseous fuel mixtures, were carried out at 20 MPa to ensure that
the trends were not being influenced by this parameter.
[0049] The effects of mixing 10% and 23% (by volume) hydrogen in
methane on emissions are shown in FIG. 3. Compared to the data from
the same engine fuelled with natural gas alone, the data from the
tests using a gaseous fuel mixture comprising 10% hydrogen showed
that for the injection timings tested, the measured emissions were
either the substantially the same or reduced. For example, the
measured data indicated that the emissions of PM, tHC and CO were
reduced on the order of 5% to 10%. Furthermore, it is noteworthy
that there were no detrimental effects to the engine operation or
the measured emission levels, resulting from the addition of
hydrogen into the fuel. That is, the addition of hydrogen had no
significant effect on the emissions of NOx.
[0050] It should be noted that the error bars presented in the
plotted data are based on the long-term uncertainty estimates,
including both analyzer sensitivity and variations in engine
operating condition. PM errors are based on calculated uncertainty
for the gravimetric samples.
[0051] The addition of 23% hydrogen had a greater impact on the
emissions than did 10% hydrogen. NOx emissions were increased
slightly but were substantially unchanged, while CO, tHC, and
CO.sub.2 (not shown) emissions were reduced. Due to uncertainties
in the PM measurements, the only observed significant influence was
at the latest timings, where a substantial reduction in PM was
observed with 23% H.sub.2 compared to the same timings for the
engine fuelled with 100% natural gas or a gaseous fuel mixture with
10% hydrogen. The presence of hydrogen in the combustion zone may
have affected pollutant emissions due to an increased concentration
of the OH radical. This highly reactive molecule would provide more
rapid oxidation of unburned fuel and partial-combustion species
such as CO and tHC. Hydrogen has also been shown to effectively
reduce local flame extinctions induced by high turbulent
strain-rates, events that are thought to generate substantial
pollutant emissions. That NOx emissions were slightly increased by
hydrogen addition is possibly due to an increase in the prompt-NO
mechanism resulting from higher OH concentrations. It may also be
due to the more intense combustion with the hydrogen addition.
[0052] The low levels of PM being measured were near the detection
limit of the instruments. However, the results shown in FIG. 3 show
that even for the gaseous fuel mixture with only 10% hydrogen a
small reduction in PM was consistently observed. For the gaseous
fuel mixture with 23% hydrogen, for earlier injection timings a
similar small reduction in PM was observed, but as the injection
timing was delayed, more significant reductions in PM emissions
were achieved. This is a significant difference in PM emissions
from what normally occurs and that is expected from engines fuelled
with 100% natural gas when later injection timings are tested.
These results show that, unlike an engine fuelled with only methane
or natural gas, by using a gaseous fuel mixture comprising methane
and at least 23% hydrogen, for a low-load, low-speed engine
condition it is possible to delay the timing for fuel injection to
achieve significant reductions in NOx emissions without the normal
consequence of significantly increasing the emissions of PM.
[0053] The effects of 10% and 23% hydrogen mixed with natural gas
are compared to the natural gas fuelling case in terms of burn
duration (10-90% of integrated heat release), gross indicated
specific fuel consumption (GISFC), peak heat-release rate, and
coefficient of variation (COV) of the GIMEP in FIG. 4. The GISFC
showed no significant influence of either timing or fuel
composition. The burn duration was substantially reduced for the
hydrogen-fuelling cases at late timing, especially with 23%
hydrogen. Interestingly, there was no change in burn duration for
the earlier timings. This suggests that different mechanisms may
restrict the combustion rate at early and late timings, with a
chemical kinetic limit at late timings, compared to a
mixing-limited condition for early timings. The peak heat-release
rate (corresponding roughly to the maximum rate of chemical energy
being released from the fuel) averaged approximately 20% higher for
the engine when fuelled with the gaseous fuel mixture comprising
23% hydrogen by volume, compared to when the engine was fuelled
with 100% natural gas. The difference when the engine was fuelled
with a gaseous fuel mixture comprising only 10% hydrogen was less
significant, although there was a slight increase in peak heat
release rate (HRR) at most timings. The use of a gaseous fuel
mixture comprising hydrogen and methane also substantially reduced
the combustion variability (as measured by the COV GIMEP). For the
gaseous fuel mixture that comprised 10% hydrogen, a significant
reduction in variability was observed at the later combustion
timings. For the gaseous fuel mixture that comprised 23% hydrogen,
reduced variability was seen at all combustion timings, although
the reduction in variability was greatest for later combustion
timings. This reduction in combustion variability can be due to
increased flame stability caused by the addition of hydrogen, which
can contribute directly to the observed reduction in CO and tHC
emissions.
[0054] FIG. 5 shows two bar charts that plot pilot and gaseous fuel
ignition delay for engines fuelled with different gaseous fuel
mixtures. Again, the data was collected from an engine operating
with the same experimental test conditions: 800 RPM; 6 bar GIMEP;
0.5 .phi.; 40% EGR, and an injection pressure of 16 MPa. As shown
in FIG. 5, for the experimental test condition, the addition of
hydrogen to natural gas had no significant effect on the pilot
ignition delay. When the engine was fuelled with a gaseous fuel
mixture comprising 10% hydrogen and 90% natural gas, there was a
slight reduction in the gas ignition delay. An on-average 20%
reduction in gas ignition delay was observed when the engine was
fuelled with a gaseous fuel mixture comprising 23% hydrogen. These
delays are defined as the time between the commanded start of
injection and the observed start of combustion. As such, they
include any physical delay within the injector, as well as both
mixing and chemical delay times for the injected fuel. The
commanded start-of-injection was a recorded value while the
start-of-combustion timing was determined by examination of the
heat-release rate. The start of pilot combustion was identified as
the first significant increase in energy release. The uncertainty
in these plots was estimated at .+-.0.5 crank angle degrees
(.degree. CA), representing the uncertainty in the crank-angle
encoder. The start of gas combustion was located as the point at
which a rapid farther increase in heat-release rate was observed.
Examples of these locations are shown in the heat-release plot in
FIG. 6.
[0055] The observed shorter gas ignition delay time is consistent
with premixed and non-premixed auto-ignition of methane tests,
previously reported in 1997 by C. G. Fotache, T. G. Kreutz and C.
K. Law in "Ignition of Hydrogen-Enriched Methane by Heated Air",
published in Combustion and Flame, Vol. 110, pp. 429-440, which
showed that hydrogen addition could substantially reduce ignition
delay times. However, the work of Fotache et al. does not relate to
a non-premixed jet being ignited by a pilot flame, and therefore is
not directly comparable to the presently disclosed method and
apparatus. Contrary to the work of Fotache et al. that suggested
that even at 10% H.sub.2, a noticeable reduction in ignition delay
occurred, the experimental data shown in FIG. 5 indicates that for
the subject internal combustion engine, which employed pilot fuel
to assist with ignition of the directly injected main fuel, a more
substantial quantity of hydrogen was required before a significant
effect was detected. Because the combustion process is complex, the
shorter gas ignition delay can have a number of effects on the
combustion process. First, the time available for mixing is
substantially reduced. While hydrogen can mix somewhat faster, due
to its higher diffusivity, the methane diffusion rate is
essentially constant. This can lead to less methane being
over-mixed during the ignition delay period, resulting in a
reduction in tHC emissions. The shorter ignition delay can also
result in less air mixing into the gaseous jet during the
pre-combustion period, resulting in a richer jet core during the
combustion process. This richer jet can result in an increase in
soot formation. The reduction in PM (which is not as substantial as
the reductions in CO and tHC) may be a result of increases in both
the soot formation (caused by the richer non-premixed jet) and
oxidation through the OH radical processes.
[0056] FIG. 6 shows that for an engine fuelled with a gaseous fuel
comprising 10% hydrogen and 90% natural gas, there was no
significant difference observed in the in-cylinder conditions, as
represented by the pressure trace and heat-release rate compared to
when the engine was fuelled with 100% natural gas. In this example,
the timing shown by 601 is when the pilot fuel injection begins,
while 602 shows the timing for when the injection of the gaseous
fuel mixture begins. The first increase in net heat release rate at
the timing shown by 603 indicates the start of combustion for the
pilot fuel and the second increase in the net heat release rate
shown by 604 indicates the timing for start of combustion for the
gaseous fuel mixture. While the pressure traces and heat-release
rates for the 0, 5, and 15.degree. ATDC timings are not shown,
similar results were observed at these other timings. When the
engine was fuelled with a gaseous fuel mixture comprising 23%
hydrogen and 77% natural gas a more significant effect on the
in-cylinder conditions was observed. For the data plotted in FIG.
7, to maintain the same combustion timing for both of the plotted
fuelling conditions (100% natural gas and a gaseous fuel mixture
comprising 23% hydrogen and 77% natural gas), to compensate for the
shorter ignition delay the timing for injecting the gaseous fuel
mixture was delayed by about 4 crank angle degrees. FIG. 7 shows
that for an engine fuelled with a gaseous fuel mixture comprising
hydrogen and natural gas, the heat release rate changes as a
function of both fuel composition and fuel injection timing. That
is, the peak heat-release rate was substantially higher at all the
combustion timings when the engine was fuelled with a gaseous fuel
mixture comprising hydrogen, with peak heat-release rate increasing
with increasing proportions of hydrogen in the fuel mixture. The
effect of peak heat-release rates being higher for engines fuelled
with fuel mixtures comprising hydrogen was relatively consistent,
although the increase in heat release rate is more substantial at
15.degree. ATDC than at the earlier timings. The effect of fuel
injection timing was observed to be consistent for both natural gas
and gaseous fuel mixtures of hydrogen and natural gas, in that
retarding injection timing resulted in reductions in the heat
release rate.
[0057] For the bulk of the testing, the mid-point of the heat
release (50% IHR) was held constant by varying the
start-of-injection timing (both pilot and main fuel timings shifted
equivalently, as the relative delay between the gas and diesel
injections was held constant). While this technique maximized
comparability of the combustion timing, it resulted in variations
in the combustion timing. To study this, experiments were conducted
to collect two sets of data. One set of data was collected from the
engine when it was operated with the same start-of-injection timing
(pilot and gas) as for when the engine was fuelled with 100%
natural gas, except that the engine was fuelled with a gaseous fuel
mixture comprising 23% hydrogen and 77% natural gas. A second set
of data was collected with the same fuelling condition but with
adjustments to the timing for start-of-injection to maintain a
constant combustion timing for the mid-point of the integrated heat
release.
[0058] The effects of these timing adjustments on the in-cylinder
performance are shown in FIG. 8, which shows the in-cylinder
pressure and heat-release rate for the following three conditions:
(1) 100% natural gas; (2) a gaseous fuel mixture comprising 23%
hydrogen and 77% natural gas, using the same timing as for 100%
natural gas; and, (3) a gaseous fuel mixture comprising 23%
hydrogen and 77% natural gas, but with the timing for
start-of-injection adjusted to maintain the same timing for the
mid-point of the integrated heat release (50% IHR) as for 100%
natural gas. This data is for the condition where the 50% IHR was
set to 10 crank angle degrees after top dead center (.degree.
ATDC), for the engine fuelled with 100% natural gas and for the
engine fuelled with 23% hydrogen with the adjusted timing for
start-of-injection. The addition of hydrogen to the fuel
substantially reduced the gas ignition delay time, as shown by the
significantly earlier main combustion event, while the pilot
start-of-combustion (shown by the first increase on the
heat-release plot) was substantially constant for all three
conditions. Similar results were seen at all timings for the
mid-point of the integrated heat release, as shown in FIG. 9. A
shorter gas ignition delay was observed for the engine when it was
fuelled with a gaseous fuel mixture comprising hydrogen under both
fixed and adjusted timings. It is thought that the gas ignition
delay was shorter for the fixed timing condition because the
ignition was occurring earlier in the cycle. The mid-point for the
integrated heat release was advanced by approximately 4 crank angle
degrees (.degree. CA) for all the constant injection timing cases.
The effects on emissions (not shown) were consistent with the
effects of advancing the timing by approximately 4.degree. CA.
[0059] The injection pressure of 16 MPa that was used to collect
most of the experimental data is lower than what is normally used
for gaseous-fuelled engines that directly inject gaseous fuels such
as natural gas into the combustion chamber of an internal
combustion engine. Generally, higher injection pressures are
considered to be more desirable and injection pressures between 19
MPa and 30 MPa are more typical. To test the effect of injection
pressure on the observed results, some of the experiments were
repeated with a fuel injection pressure at 20 MPa. While this was
still substantially below the highest achievable injection
pressures, it provided a reasonable injection rail/peak cylinder
pressure ratio, due to the low in-cylinder pressure. The minimum
fuel/cylinder pressure ratio at the earliest combustion timing
(where the peak cylinder pressure was highest) was 2:1 at 20 MPa,
compared to 1.6:1 for the 16 MPa injection. For later combustion
timings, the ratio was increased to as much as 3.3 (compared to 2.7
for the 16 MPa case). These ratios do not represent the actual
ratio between the fuel at the injector nozzle and the in-cylinder
condition, as the cylinder pressure changed over the injection
period, while the pressure of the gas exiting the nozzle was
substantially lower than the rail pressure due to flow losses
within the injector body and gas dynamics at the nozzle outlet.
However, these ratios do provide a means for characterizing the
effect of injection pressure and provide a basis for comparing such
effects between engines that are fuelled with 100% natural gas, and
engines that are fuelled with a gaseous fuel mixture comprising
hydrogen and methane.
[0060] The effect of increasing the injection pressure on emissions
is shown in FIG. 10. The higher injection pressure tended to
increase CO, PM, and tHC emissions, while NOx and GISFC were not
affected. The results can be seen to be consistent for both natural
gas and hydrogen-methane blend fuelling. That the injection
pressure had little impact on the in-cylinder performance is shown
in FIG. 11, which plots a pressure trace and the heat release rate
for 16 and 20 MPa, for an engine fuelled with a gaseous fuel
mixture comprising 23% hydrogen and 77% natural gas by volume. It
was surprising that the higher injection pressure resulted in
slightly increased levels of PM, tHC, and CO emissions compared to
engine fuelled with the same gaseous fuel mixture but with lower
injection pressures. However, the experimental results do show that
hydrogen addition resulted in reductions in the emissions of PM,
total hydrocarbons (tHC) and carbon monoxide (CO), without negative
impact on emissions of NOx, and that this result was generally
consistent at both injection pressures. Accordingly, these results
indicate that hydrogen addition has a positive impact on emissions
over a range of fuel injection pressures.
[0061] From the experimental data collected it is possible to
determine certain trends relating to engine emissions and
combustion stability arising from fuel composition and combustion
timing. That is, these trends can be extrapolated from the data
that was collected when the engine was operated with gaseous fuel
mixtures comprising 100% natural gas (and 0% hydrogen), 90% natural
gas and 10% hydrogen, and 77% natural gas and 23% hydrogen. When
the engine was fuelled with 90% natural gas and 10% hydrogen,
improvements were observed in combustion stability and engine
emissions were substantially the same or slightly reduced compared
to when the same engine was fuelled with 100% natural gas. When the
same engine was fuelled with 77% natural gas and 23% hydrogen,
there were greater improvements in combustion stability and more
substantial improvements in engine emissions. Although the results
are not plotted in the figures, experiments were also conducted in
which the engine was fueled with up to 35% hydrogen by volume (at
STP), and at such higher hydrogen percentages the effect on
emissions continued to be beneficial. However, hydrogen is harder
to compress compared to natural gas and the higher volume occupied
by hydrogen compared to methane for the same amount of energy
introduces volumetric flow capacity challenges for gaseous fuel
mixtures with higher percentages of hydrogen. From the experimental
data collected, the levels of emissions observed from the conducted
experiments, and the pre-existing knowledge base relating to the
combustion of gaseous fuel mixtures in other engines, it can be
reasonably determined that, compared to an engine fuelled with 100%
natural gas, improved combustion stability and improved engine
emissions can be achieved with gaseous fuel mixtures comprising
hydrogen in concentrations from 5% to at least 60% by volume. From
the observed trends plotted in FIGS. 3 and 4, higher hydrogen
concentrations can yield better combustion stability (reduced
combustion variability) and lower emissions, but these advantages
can be offset by other factors such as higher hydrogen percentages
requiring increased volumetric flow requirements, or the cost and
availability of hydrogen. For higher percentages of hydrogen, the
properties of the gaseous fuel mixture can also change because
hydrogen has a lower lubricity compared to natural gas. In some
cases, the preferred gaseous fuel mixture can be between 10% and
50% hydrogen or an even narrower ranges, such as between 15% and
40% hydrogen mixed with natural gas or between 20% and 35% hydrogen
mixed with natural gas. By way of specific examples, the gaseous
fuel mixture can comprise methane and hydrogen with hydrogen
content expressed as a percentage by volume being one of 12%, 14%,
16%, 18%, 20%, 22%, 23%, 24%, 25%, 26%, 28%, 30%, 32%, 34%, 35%,
36%, 38%, 40%, 42%, 44%, 46%, 48%, 50% and percentages
therebetween.
[0062] From the experimental data, trends can also be determined
relating to combustion timing. For an engine fuelled with a gaseous
fuel mixture comprising hydrogen and methane, combustion stability
can be achieved over a broader range compared to the same engine
fuelled with 100% natural gas. For an engine fuelled with a gaseous
fuel mixture comprising 10% hydrogen, this improved stability was
observed to occur when timing for the mid-point of the integrated
heat release occurred 10 crank angle degrees after top dead center
and later. For the same engine fuelled with a gaseous fuel mixture
comprising 23% hydrogen, an improvement in combustion stability was
observed as early as when the mid-point of the integrated heat
release occurred 5 crank angle degrees after top dead center, with
improvements to combustion stability increasing further still for
later combustion timings. From the experimental data it can be
concluded that an engine fuelled with a gaseous fuel mixture
comprising methane and at least 10% hydrogen by volume, can equal
or better the combustion stability and emissions from the same
engine fuelled with 100% natural gas. Even though most of the data
was collected for one engine operating condition, since the
selected engine operating condition was one that is normally
associated with high engine emissions it is expected that the
tested gaseous fuel mixtures comprising at least 10% hydrogen and a
majority of methane by volume will produce similar or better
emissions and combustion stability compared to the same engine
fuelled with 100% natural gas, when the engine is operated at
different engine conditions.
[0063] In summary, the experimental results show that an internal
combustion engine with direct injection of a gaseous fuel mixture
comprising hydrogen and methane can be operated to reduce emissions
and improve combustion stability compared to the same engine
fuelled with 100% natural gas. The graph of coefficient of
variation of the GIMEP against combustion timing in FIG. 4 shows
that the addition of hydrogen results in a substantial reduction in
the combustion variability. The experimental results also show that
while hydrogen addition can increase the peak combustion heat
release rate, indicating higher combustion temperatures, the
addition of hydrogen did not result in increased levels of NOx
emissions compared to when the engine was operated under the same
conditions but fuelled with 100% natural gas. The results further
show that hydrogen addition can allow later combustion timings
because the level of PM emissions at later combustion timings are
reduced compared to when the engine was fuelled with 100% natural
gas. The experimental data confirmed that like engines fuelled with
100% natural gas, the levels of NOx emissions decrease with later
combustion timings for engines fuelled with gaseous mixtures
comprising hydrogen and methane. Whereas with engines fuelled with
100% natural gas, the steep increase in PM emissions for later
combustion timings establishes a limit to how much combustion
timing can be retarded, the experimental results show that for
engines fuelled with a gaseous fuel mixture comprising hydrogen and
natural gas, later combustion timings are possible because PM
emissions increase at a much shallower slope as combustion timing
is delayed. In addition, it was found that a characteristic of
gaseous fuel mixtures comprising hydrogen and methane that were
directly injected into a combustion chamber of an internal
combustion engine was that the gaseous fuel mixtures ignited with a
shorter ignition delay compared to that of natural gas without the
addition of hydrogen. For the tested engine condition the shorter
ignition delay results in the combustion timing being advanced
about 4 crank angle degrees, which resulted in higher peak
in-cylinder pressures and higher peak heat release rates if the
same injection timing used for a natural gas engine was maintained.
It was determined that timing adjustments can be made so that
combustion characteristics match those of engines fuelled with 100%
natural gas.
[0064] While particular elements, embodiments and applications of
the present invention have been shown and described, it will be
understood, that the invention is not limited thereto since
modifications can be made by those skilled in the art without
departing from the scope of the present disclosure, particularly in
light of the foregoing teachings.
* * * * *